FIG. 6 is a sectional view along a rotational axis line illustrating a conventional example of a radial-flow type exhaust turbo-charger with the aforesaid radial compressor built therein.
Referring to FIG. 6, reference numeral 10 denotes a turbine casing and reference numeral 11 denotes a scroll formed spirally around the outer periphery of the turbine casing 10. Reference numeral 12 denotes a radial-flow type turbine rotor provided coaxially with an impeller 8, and a turbine shaft 12a thereof is rotatively supported by a bearing housing 13 through the intermediary of a bearing 16.
Reference numeral 7 denotes a compressor housing which accommodates the impeller 8, reference numeral 9 denotes an air inlet passage of the compressor housing 7, and reference numeral 7a denotes a spiral air passage. Reference numeral 4 denotes a diffuser. These components constitute a radial compressor 100. Further, reference numeral 100a denotes a rotational axis center of the exhaust turbo-charger.
When the exhaust turbo-charger constituted as described above operates, an exhaust gas from an engine (not shown) enters the scroll 11, flows from the scroll 11 into a turbine rotor 12 from the outer periphery side thereof, and flows in a radial direction toward a central side to impart dilatational work on the turbine rotor 12. Thereafter, the exhaust gas flows out in the axial direction and is sent out of the exhaust turbo-charger by being guided to a gas outlet 10a. 
The rotation of the turbine rotor 12 causes the impeller 8 of the radial compressor 100 to rotate through the intermediary of the turbine shaft 12a. The air taken in through the air inlet passage 9 of the compressor housing 7 is pressurized by the impeller 8, and then the pressurized air is supplied to the engine (not shown) through the air passage 7a. 
The radial compressor 100 of the exhaust turbo-charger described above can be stably operated according to a relationship between a choke flow rate and a surge flow rate of air, as illustrated in FIG. 10(B). However, the range of flow rate permitting the stable operation is limited, so that it is necessary to operate the radial compressor 100 at a low-efficiency operating point away from a surge flow rate so as not to induce surging during a transient change at a rapid acceleration.
The radial compressor 100 presents a significant drawback in that the flow rate range between the choke flow rate and the surge flow rate becomes narrow, as illustrated in FIG. 10(B), due to the occurrence of the surging.
The surging is caused by a stall of a flow at an inlet of the impeller 8 or by a stall of the diffuser 4.
The flow at the inlet of the impeller 8 of the radial compressor 100 changes with flow rate. As illustrated in FIG. 10(B), the stable operation is performed according to the relationship between the choke flow rate and the surge flow rate; however, the stable operation cannot be performed at a flow rate of the surge flow rate or less.
At a normal operating point, as illustrated in FIG. 10(C1), a flow smoothly comes in between blades 8a of the impeller 8 along the contours of the front ends of the blades 8a of the impeller 8. However, at the surge flow rate, a stall 9a′ of the flow at the front ends of the blades 8a takes place, as illustrated in FIG. 10(C2). The stall 9a′ of the flow at the front ends of the blades 8a of the impeller 8 is one of the causes of the occurrence of surging.
The occurrence of surging is generally attributable to the stall 9a′ in the impeller 8 or the stall of the diffuser 4. The present invention is focused mainly on the improvement of the surging (a reduction in a surge flow rate) attributable to the impeller 8.
As a means for preventing the occurrence of the surging, there has been one proposed in Patent Document 1 (Japanese Patent Application Laid-Open No. 58-18600).
FIGS. 8(A), (B), and (C) illustrate flows in the vicinity of surging which has occurred in the current impeller 8. As the flow rate reduces due to a stall at the inlet of the blade 8a of the impeller 8, an incidence angle w of the flow increases and a flow 9f begins to come in from an upstream of the blade 8a toward a pressure plane, as illustrated in FIG. 8(B). This flow leads to the occurrence of the so-called stall phenomenon in which the flow 9f breaks away on a negative pressure plane when the aforesaid flow turns in to the front end of the blade 8a (a backflow takes place on the negative pressure plane).
The stall phenomenon at the blade 8a causes a further increase in the incidence angle w of a flow coming to a blade 8a′, which is on the reverse rotation side from the blade 8a, resulting in larger separation on the blade 8a′. This phenomenon is propagated to the blade 8a′ on the reverse rotation side and a backflow 9g occurs also on a negative pressure plane by a backflow 9h reaching the negative pressure plane from a pressure plane 8a1 beyond the front end of the blade 8a, as illustrated in FIG. 8(C).
Thus, the stall phenomenon of the impeller 8 expands with a consequent pressure drop of the impeller 8, and surging takes place.
As a means for preventing the occurrence of the surging, there has been one proposed in Patent Document 1 (Japanese Patent Application Laid-Open No. 58-18600). In the means, as illustrated in FIGS. 9(A) and (B), an annular concave groove 7b is formed in the peripheral wall of the air inlet passage 9 of the compressor housing 7, and a rear end portion of an opening of the annular concave groove 7b which meets a housing peripheral wall 3 of the annular concave groove 7b is provided such that the rear end portion extends over a blade front end surface 1 of the impeller 8. The rear end portion of the opening of the annular concave groove 7b is provided at a downstream of the front end surface of the impeller so as to allow a circulating flow 18′ to pass by the distal end of the impeller between the front end surface of the impeller and the rear end of the impeller.
In this case, as illustrated in FIG. 9(A), in the case where the rear end portion of the opening of the annular concave groove 7b is provided so as to extend over the blade front end surface 1 of the impeller 8, and the radius of the housing peripheral wall 3 of the air inlet passage 9 agrees with the radius of a peripheral wall 3′ of a casing at the outlet side of the annular concave groove 7b, a backflow vortex 18′ passing by the blade distal end at the downstream of the blade front end surface occurs due to a centrifugal force in a small-flow-rate area.
Further, as illustrated in FIG. 9(B) (FIG. 17 in Patent Document 1), providing the rear end portion of the opening of the annular concave groove 7b such that it extends over the blade front end surface 1 of the impeller 8 and setting the radius of the housing peripheral wall 3 of the air inlet passage 9 of the annular concave groove to be larger by U than the radius of the peripheral wall 3′ of the casing on the outlet side balances a centrifugal force and the dynamic pressure on the upstream side by a design flow rate. This ensures smooth flow of a mainstream.
In this case, the rear end portion of the opening of the annular concave groove 7b is provided such that it extends over the blade front end surface 1 of the impeller 8. A relationship is illustrated that the blade front end surface 1 of the impeller 8 extends over the rear end portion of the opening of the annular concave groove 7b, and the blade distal end portion is configured so as to allow a circulating flow to pass thereby. This poses a drawback in that performance deteriorates at a normal operating point.